Negative spring compensation for elastomeric bearing torque

ABSTRACT

Elastomeric bearings exhibit a reactive positive spring effect, i.e., they produce an opposing torque or force when angular displacement or translation is applied upon them, due to shear stress developed within their elastomer parts. The present invention incorporates a device that exhibits an increasingly strong torque or force in the same direction as the displacement, i.e., a negative spring, under similar conditions of movement. When properly calibrated, the subject type of device can be used in paralleled motion with the ordinary reactive elastomeric bearing to produce a combined effect in which at least a part of the torque or force of the elastomeric bearing is changed or minimized over a range of movement.

BACKGROUND OF THE INVENTION

The field of the invention relates to rubber or elastomeric bearings(including laminated rubber bearings) used to support limited-movementbetween opposing loading members, all of which develop torques or forcesmore or less proportional to the extent of movement between the opposingmembers over a range, essentially an elastomeric spring effect. In somecases, the torques or forces required may exceed those ordinarilyavailable, as exerted by humans for instance, and powered boosters havebeen necessary to achieve the desired extent of movement.

The following United States of America Patents are cited as references:

2,900,182 Hinks 3,228,673 Hinks 3,532,174 Diamantides et al 3,734,546Herbert, et al 3,504,902 Irwin 6,524,007 Hinks 6,834,998 Hinks 5,794,753Kemper 5,887,691 Kemper 5,967,283 Kemper 5,984,071 Kemper 4,607,382Dijkstra 4,722,517 Dijkstra 5,178,357 Platus 5,310,157 Platus 5,390,892Platus 5,669,594 Platus

The following Canadian Patent is a further reference:

731007 Ballauer

In the prior art, elastomeric bearings as disclosed in U.S. Pat. Nos.2,900,182 and 3,228,673 include at least one, but usually multiplealternate laminations of metal or other strong inextensible material andrubber or elastomer usually bonded together. Lateral motions betweensucceeding metal laminations are permitted by shear strain within andparallel to the intervening rubber laminations. They can be made withlayers in any shape, with apertures or not, and with variouscross-sectional configurations, including truncated planar, conical,spherical, chevron-shaped or cylindrical layers.

All elastomeric bearings are used to separate and support opposedrelatively moveable external loading members that bear upon the outerload-accepting layers or end pieces of the bearings that have load facesand are generally made of thicker metal. The opposing outer layers maybe shaped to conform with and to seal with respect to their respectivecomplemental loading members and to provide for keying to the latter fororientation and prevention of relative slipping.

When the external load faces of such a bearing are interposed betweensuch complementally-contoured and opposed loading members, it can resistthrust, radial or combined forces normal to its layers, depending uponits configuration. Relative lateral movement between the opposed loadingmembers, which may include pivoting about a normal axis as well astransverse or lateral shifting, results in a distribution of shearingmovements between individual rubber layers.

An additional property of such a load-bearing bonded laminate stack thatcontains one or more apertures is the capability of preventing thelateral or transverse flow of fluids, i.e., liquids or gases, betweenthe periphery of the laminate stack and an aperture, and making themessentially impervious even under substantial differential pressure.I.e., the space occupied by the bulk of the laminations between theopposing members is blocked against fluid penetration. U.S. Pat. Nos.3,532,174, 3,734,546, 3,504,902, 6,524,007 and 6,834,998 exhibit theconcept of rubber laminated bearings that seal against fluid flow,referred to here as bearing-seals. This sealing property is neverthelessirrelevant to the current invention.

As indicated above, elastomeric bearings and bearing-seals usually havethe primary purpose of supporting loads and/or sealing between opposingmembers while permitting limited motion between said members, whetherrotational or translational. Since that motion is the cumulative resultof shear strain in the layers of elastomer itself, these devices usuallydevelop negligible coulomb friction, but do exhibit an increasingresistive or reactive force or torque due to shear stress in theelastomer layers that accompanies the motion. This essentially linearspring effect can be described over the effective range by a numberrepresenting the rate of change of reactive force or torque actingagainst the displacement, i.e., its translational or torsionalstiffness, i.e., spring rate.

In some cases, this stiffness is negligible in comparison to the forcesor torques available to overcome them, and in others, it is a desirableeffect. However, in situations where the high reaction forces or torquesof elastomeric bearings exceed those of the means readily available tocounteract them, those means have often been replaced, amplified, orsupplemented by power booster means, which may be complex, expensive,unreliable, and weighty or otherwise undesirable.

This has often been particularly true for helicopter control systems.Laminated elastomeric bearings have frequently been made part ofhelicopter rotor hubs to retain each of the rotor blades against veryhigh centrifugal forces while permitting their blade pitch, i.e.,feathering, angles to be changed for control purposes. But except forsmall helicopters, it has been found that the forces required to changethe pitch of the elastomeric-retained blades generally exceeds thoseavailable through human actuation of the pitch control sticks alone, andhydraulic boosters have conventionally been used to relieve the pilotfrom high control stick forces.

A similar situation, in principle, was faced by Kemper (U.S. Pat. No.5,794,753, etc.) in a problem associated with the human-operated clutchof heavy trucks and other machinery. But rather than rely onconventional externally-powered actuators to help operate the clutch, hedescribes systems involving passive Bellville springs to accomplish thatpurpose. Bellville springs possess a non-linear force-displacementbehavior that includes a region in which extended motion causes not aproportionally resisting force, but instead a force in the samedirection and increasing with the motion. This is in effect defines anegative spring rate region of their force-displacement characteristic.Thus the Bellville spring's proportionally increasing force to separatethe clutch plates helps the human act against the clutch springs thatclamp the rotating clutch plates together.

Dijkstra (U.S. Pat. Nos. 4,607,382 and 4,722,517) and others referencedtherein employ negative spring means to reduce the effective stiffnessof loudspeaker cones and thereby lower their natural frequency.

Platus (U.S. Pat. No. 5,178,357, etc.) describes vibration-isolationplatforms that employ a relatively stiff spring to support the weight ofa mass placed upon the platform, together with negative spring meansacting in parallel to reduce the effective local stiffness of thecombined springs. This reduces the resonant frequency of the suspendedspring-mass system so that results of higher frequency test vibrationsapplied to the mass are essentially unaffected by the suspension system.

Besides the inventions of Kemper, Dijkstra and Platus, many commonextant devices employ elements that produce some characteristics ofnegative springs through involvement of sources of pushing or pullingforces. These force sources include passive springs of various kinds,such as coil, leaf, Bellville and Neg'ator springs, used in eithertension or compression, and actuators powered by hydraulics, pneumatics,or electromagnetics, etc. The common quality of all these negativespring devices is that, with respect to some “center” position, they allexhibit a characteristic torque or force versus deflection behavior ofurging movement farther away from the center over a range when initiallydeflected away from it, i.e., they exhibit a binary instability or“over-center” effect.

For instance, the common “snap-action” electrical switch, having manyvariations, often uses a pivoted compression spring that tends to forcethe associated contact assembly into one of two stable positions, eitherthe “ON” or “OFF” state. When the spring is moved by a switch handlethrough its tightly compressed center position and goes over-center, theforce of the spring on the contact assembly changes direction abruptlyand causes it to quickly change states. A motion away from center causesa component of force to develop urging further movement in the samedirection away from center. But in this case, there is no attempt to usethis repelling effect to quantitatively compensate any positive springcontinuously over a range of motion, but only to cause motion of thecontact assembly as far as it will go in either direction.

On the other hand, Kemper and Dijkstra and Platus do calibrate theirnegative spring functionalities against the primary positive springeffects of their devices. Of these inventors, the first apparently usesthe over-center effect only on one side of the force-centered position,while the others' apparatus operates on both sides of center.

SUMMARY OF THE INVENTION

The purpose of the invention is to provide a simple, passive,inexpensive, and lightweight means of overcoming undesirable highstiffness characteristics of elastomeric bearings and/or seals, whetherthey are rotational or translational types, responding respectively toapplied torques or forces. This purpose applies especially forhelicopters,

Much of the following description applies to both types whether the term“torque” or “force” is used.

Although elastomeric bearings are the focus of discussion throughoutthis disclosure, there exist other mechanisms that also exhibit the kindof springlike characteristics that elastomeric bearings have. Theseinclude so-called tension-torsion straps as used to retain helicopterrotor blades, and certain metallic joints (e.g., C-Flex), both of whichhave torsional spring properties. Most of the following discussion andclaims are applicable to these types of bearing and support devices aswell.

In brief, the invention involves the association of negative springmeans together with elastomeric bearings and/or seals so that saidnegative spring means experience at least part of the same orproportional motion as, and produce a combined result of reducing thepositive reaction torques of, said elastomeric bearings and/or seals. Toaccomplish these results, the negative spring means must exhibit atorque or force that acts in the direction away from a zero force centerposition, i.e., as if repelled from said center position. This is anessentially linear effect that can be described over its effective rangeby a negative number representing the rate of change of the negativespring means' force or torque with deflection. In some cases thisnegative rate would numerically be equal in magnitude as well asopposite in sense to the positive rate of the elastomeric spring effect,thereby canceling it and rendering the overall change of force or torquenegligible over a range of motion.

It is an object of the invention to employ negative spring means toreduce or modify the force or torque required to operate elastomericbearings and/or seals through at least part of their movement regime.

BRIEF DESCRIPTION OF THE DRAWINGS

FIGS. 1, 2 and 3 are graphs showing force or torque vs. deflectioncurves of elastomeric bearings and negative springs according to theinvention.

FIG. 4 schematically represents a simple type of mechanism includingpushing means that functions as a negative spring according to theinvention.

FIGS. 5 and 6 show details of two types of compression springs thatcould be used for a force source.

FIG. 7 shows a negative spring device similar to FIG. 4 used to directlycancel lateral spring forces developed by a translational elastomericbearing according to the invention.

FIGS. 8 and 9 are a side view and a sectional view of a device thatshows use of compressed leaf springs in a negative spring device pairedwith a radial-loadable elastomeric bearing according to the invention.

FIG. 10 shows use of an elastic compressive shell in an angular-actingnegative spring device according to the invention.

FIG. 11 shows a side cross-sectional view of negative spring meansclose-coupled with a spherical elastomeric bearing that is compensatedby it according to the invention.

FIGS. 12 and 13 are top and partly cut-away side views of a negativespring device paired with an elastomeric thrust bearing according to theinvention.

FIGS. 14 and 15 are top and partly cut-away side views of a negativespring device paired with an elastomeric thrust bearing havinglaminations that are segments of spheres according to the invention.

FIGS. 16 and 17 are top and side cross-sectional views respectively ofanother configuration of rotational negative spring device according tothe invention.

FIGS. 18 through 20 show several versions of levers plus push or pulldevices that implement negative springs according to the invention.

FIG. 21 shows the application of the negative spring device of FIG. 18to the collective pitch stick of a helicopter.

FIG. 22 indicates the attachment of the negative spring device accordingto FIGS. 16 and 17 to the collective pitch stick of a helicopter.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 illustrates the theoretical force or torque vs. deflection curvesof elastomeric bearings, negative springs, and the combined result ofthese opposing torques or forces at any given deflection. The horizontalaxis represents the angular or lineal deflection d, negative at the leftand positive at right relative to zero deflection at the center, whilethe vertical axis represents units of torque t or force f, positive upand negative down relative to the central zero.

The point of zero deflection and zero torque (or force) represents therelaxed, untorqued state of the elastomeric bearing (conveniently placedat the origin O of the plot), and the straight line e through this pointis the idealized torque response of the bearing as it is rotated througha deflection angle d. It has the positive slope Se (i.e., a positivespring rate or stiffness), meaning that the torque developed opposes thedirection of deflection, as if to restore it back to the zero position.In the form of an equation Se=Dte/Dde, where Dte=change of elastomerictorque and Dde=change of deflection of the elastomeric spring. Thistorque response is idealized as a straight line, although the realbearing will generally deviate substantially from linearity at largevalues of deflection as the elastomer reaches the limits of itsresilience and generally becomes stiffer, as indicated by the dashedlines labeled e′.

Similarly, an ideal t vs. d plot of the negative spring labeled n isseen, depicting a force or torque equal and opposite to that of theelastomeric spring for every value of deflection, i.e., this torque actsin the same direction as the deflection, so as to aid it. The straightline slope Sn (i.e., negative spring rate) is represented by theequation Sn=Dtn/Ddn, where Dtn=change of negative spring torque andDdn=change of negative spring deflection. By definition, Sn=−Se, Ofcourse, this curve will likely deviate somewhat from linearity also,particularly near the limits of its working range as indicated by thedashed lines labeled n′.

In this ideal case the negative spring torque (force) exactly cancelsthe elastomer torque at every point, and the resulting combined force ortorque is exactly zero for all deflections in the linear range. Hence,the combined t vs. d curve lies on the horizontal axis, labeled c. Thedashed deviation lines labeled c′ represent the combined actualdeviations from true cancellation at the limits of the range ofdeflection. Under these conditions and within the effective range, therewould theoretically be little or no steady force exerted by a helicopterpilot on his control stick, and it would remain in place if he releasedit from his hand.

The case of exact cancellation shown in FIG. 1 requires both theelastomeric bearing and the negative spring means to have identicalabsolute spring rates (though in the opposite sense) and for both tohave their center locations (deflections with zero force) aligned. Thatsituation may not be the case, either by inexact compensation or bydesign.

FIG. 2 is another example of an ideal t vs. d plot of the positive andnegative springs over their active range with the same designations. Theelastomer spring and negative spring plots are again arbitrarily assumedto have torque centers that coincide at a point, which is taken to bethe origin of the graph as in FIG. 1. However, the negative slope issmaller in absolute magnitude than the positive spring slope. Thecombined plot (obtained by subtracting the torque magnitude of thenegative spring from the positive torque at all deflections) is seen bythe dashed line to exhibit a smaller spring rate c (i.e., less stiff)than the original elastomeric bearing stiffness. This might have adesirable effect of making human control forces manageable while stillcausing a control stick to revert to the neutral center position withhands off. There may be applications where making the negative springstronger, with the attendant binary instability would be beneficial,

FIG. 3 is yet another example. The elastomeric spring plot is assumed tohave its torque center at the origin of the graph, but the torque centerof the negative spring is offset by the distance labeled dO, while thenegative slope has the same absolute magnitude as the positive springslope or spring rate. The combined plot is seen to be a constant smalltorque tO in one direction, depending upon the relative registration orlocations of the force centers. It can be seen that numerically,tO=dO*Sn.

In general, combinations of the conditions of FIGS. 2 and 3 may havebenefits in various applications of elastomeric bearings.

FIG. 4 schematically represents one simple type of negative springmechanism that employs an extensible pushing or repelling force means(termed a “pusher” here), such as a coil or bent leaf spring, a gascompression spring, or a hydraulic or electromagnetic source of force F.The generic pusher 3, indicated by heavy dashed lines, has ends 1 and 2,and force F acts against each along the straight line joining them.i.e., the chord. Both said ends have moveable or pivoting end joints:pin or knife-edge ends captured by complemental V-shaped receptacles areshown in this example. In other implementations, the end pivots may behinged joints or flexible metal linkages.

End 1 is captured by a frame member 7 and end 2 bears upon a member 4moveable with respect to member 7 such that end 2 is constrained to movealong a path 6 that has a center position 5 where the chord line offorce 1-2 is perpendicular to the path line 6. Means permitting movementof member 4 along path 6 is schematically depicted by but not restrictedto rollers 8, seen in end view.

At said center position 5, pusher 3 has maximum compression, i.e., theshortest length ds of its dashed-line chord 30 extending from end 1 toend 2 of the pusher at position 5. In the simplest form, the path 6 iseverywhere perpendicular to the chord line 30 at its center Position.

It is implicitly assumed here that the compressive force F createdwithin pusher3 between its end points 1 and 2 can be represented by avector that has two force components one of which is perpendicular topath 6 at any point, and the other (of primary interest) is aligned withpath 6. This component of force aligned with path 6 is the negativespring force fn. In center 5 position, that force fn along the path 6 iszero, but as end 2 moves with member 4 on path 6 up or down either wayfrom the center 5, the chord of the pusher 3 will take the positionsindicated by lines 31 or 32, and the lateral component of force fnparallel to path 6 will be developed by end 2 upon member 4, urging itfarther away from center 5. In fact, if the force applied by the pusher3 along its chord is F, and the distance away from center 5 along path 6is d, the force fn developed along the path 6 upon member 4 is given bythe relationship fn=F*sin[arctan (d/ds)]. If pusher 3 is an ordinaryspring, however, the force F will drop off as its length increasesbeyond the minimum ds, resulting in a more complex equation althoughstill nearly linear over a range of interest.

It may be necessary to provide end stops 33 and 34 as shown to limit theexcursions of end 2 to keep the device functional at large deflections.

Considering the system of FIG. 4 as a free body diagram, as the force fnupon member or element 4 is developed, there must be an equal andopposite force fn developed upon member or element 7, and means must beprovided at appropriate attachment points to apply these opposing forcesagainst the reactive forces developed between the two loading members ofan associated elastomeric bearing, i.e., said elements 4 and 7 of thenegative spring mechanism must respectively be mechanically connectedbeneficially to said two loading members.

FIGS. 5 and 6 show details of two types of compression springarrangements that could be used for the pusher 3 force source indicatedschematically in FIG. 4. In particular, FIG. 5 uses a wire coil spring36 that may be stabilized against columnar buckling by two encircled andopposed end parts 35, 35 that support the ID of the spring while alsoproviding pin ends. The end parts are in turn maintained co-linear by arod 37 telescoped within them. As in FIG. 4, the end supports 7 and 4accept compression force F from the spring and move laterally withrespect to each other, as along the path 6, thereby developing thelateral force fn along that path.

The compression spring of FIG. 6 is a buckled-columnar leaf or bladespring 3 of constant or variable width and thickness, shown increscent-shaped edge-on view, with knife-edge ends at 1 and 2. Thespring contours at locations 31 a and 32 a occur at the opposite maximumrelative lateral extents of movement between members 7 and 4 as the bentspring extends and becomes somewhat less bowed. However, its chordlength, the distance along the lines 31 or 32 between end points 1 and2, is maintained less than its relaxed chord length, and as well known,this buckled configuration will produce a higher force F between itsends as it straightens out. This strengthening effect may tend toincrease the quasi-linear range of useful compensation by force fncompared to a spring that weakens with extension.

Blade springs having flexible metal end joints were used advantageouslyby Dijkstra, and could be used here in place of the leaf spring withknife-edge ends as shown, whether single or doubled per Dijkstra.However, making use of end compliance to provide the necessary endrotation would introduce some amount of positive (springy) restoringeffect resulting from end-flexing, thus reducing their negative rateeffects.

FIG. 7 illustrates the principal of the negative spring means of FIG. 4,applied directly in conjunction with a translational elastomeric bearing9. As noted earlier, bearing 9 is comprised of usually multiplealternate layers or laminations of metal or other strong inextensiblematerial and rubber or elastomer (layers shown in edge view by parallellines), to define a rectangular thick pad in this case.

Bearing 9 has the primary function to support a normal load L placedupon it by the loading member 4 while backed up by loading member 7, assaid bearing is forced to move down as shown with member 4 by anexternal force E along lateral path 6 through a distance d. The movementalong path 6 by member 4 is permitted by shear strain within the rubberlaminations as cumulatively indicated by slanted dashed edge lines 10,and develops a reactive force fe that is essentially proportional todistance moved, as shown in the plots of FIGS. 1-3

Negative spring means are introduced to compensate for said reactiveforce of bearing 9 as loading member 4 moves along path 6, in this caseby arranging for pusher means 3 to act directly upon member 4 as in FIG.4. The rollers 8 of FIG. 4 can be replaced as a matter of convenience bythe laminated bearing 9 since the bearing provides the means for lateralmotion instead. However, although the forcing means 3 act in the planeof the paper in FIG. 7, the result will be the same in 3 dimensions aslong as the force F acts upon member 4 in the neutral position fromwithin a plane that includes point 1 and is normal to the path line 6and perpendicular to the paper. In any case, the negative spring unitwould share the same lateral deflection seen by said bearing, therebydeveloping the aiding force fn that directly counteracts or reduces theoverall reactive lateral force fe, in accordance with FIGS. 1-3.

The principle of FIG. 4 can also be adapted as negative spring means tocompensate directly for the reactive torque of angular rotationelastomeric bearings.

FIGS. 8 and 9 show one such approach, in which a radially-loadablelaminated rubber bearing 9 is seen in cross-section in FIG. 9. Itslaminations are shown there edge-on as closely spaced lines parallel toits central axis, and are seen in the right side broken-away part ofFIG. 8 as concentric lines. Relative angular rotation applied betweeninner ring 7 and outer ring 4, i.e., the loading members, indicated byarc 6 will result in more or less proportional reactive torque tobetween them as seen in FIGS. 1-3.

Negative spring means are implemented via widthwise extensions of therings 7 and 4 (not necessarily of the same diameters as shown). Saidrings embrace in the extended annulus between them a multiplicity ofcompressed and buckled columnar-end-loaded leaf springs or other typesof pushers 3, each of which has ends 1 and 2 that fit into notches inring 7 and 4 respectively. The pushers 3 are arranged with the centralpositions of their chords radially-oriented within the annulus. Here,instead of the lateral path being a straight line 6 as in FIG. 4, therings establish the curved path 6 for the motions of the ends 2 of eachpusher 3. For clarity, only the top one of the pushers 3 with ends 1 and2 is shown by dashed lines 31 a and 32 a to indicate angular movement ofthe outer ring relative to the inner ring through an angle a. As thatmovement occurs between said rings, a deflection-aiding torque to willdevelop between them as seen in FIGS. 1-3 to compensate against saidreactive torque.

The instability of this arrangement is obviously similar to that of FIG.4, and in the absence of the paralleled laminate bearing, if the innerring 7 were fixed, the outer ring 4 would be torqued to move angularlyabout the central axis by the tangential components of each of thein-line forces F of the pushers 3 as their chords deviate one way or theother from their neutral radial orientations. The purposes of theinvention will be achieved when the thereby-implemented negative springrates are properly calibrated relative to the positive spring rate ofthe associated elastomeric bearing according to FIGS. 1-3, and thealignment or desired relative registration of their respective torquecenters is realized.

Although FIG. 8 shows shallowly-bent leaf springs 3 nested together withtheir width dimensions more-or-less parallel to the concentric axis ofrings 4 and 7, their bent shape could be S-shaped, or bent wire springsthat extend the body of their curvature outwardly parallel to said axisfrom their radially-oriented ends 1 and 2. Close nesting of curvedsprings could result in efficient use of space for a large effect.

FIG. 10 shows radial negative spring means using an outer ring or shell4 that, instead of being rigid, itself has the qualities of ahoop-tension spring. It squeezes down upon and thereby createscompression forces in radially oriented pins or pegs 3 with rotatableends 1 and 2 as pushers in place of the previously shown radialcompression springs. The inner ring 7 could otherwise be an expandinghoop-spring that accomplishes the same purpose. Such hoop-tension orcompression spring properties of the outer or inner rings, respectively,might be realizable through incorporation of convolutions 10 or othermeans of enhancing elastic qualities in the rings, including elastomericmaterials, rather than metal.

The concept of FIG. 4 can be extended to other configurations ofelastomeric bearings, including conical and spherical as in FIGS. 3 and8 respectively of U.S. Pat. No. 2,900,182.

FIG. 11 shows a sectional view of such a spherical elastomeric bearingas compensated by negative spring means in the form of multiplecircumferentially-spaced pushers 3 having ends 1 and 2, acting upon anassociated spherical bearing 9, all contained within a supportinghousing and loading member 7.

As shown in U.S. Pat. Nos. 6,524,007 and 6,834,998, FIGS. 4 an 7, insaid spherical bearing each individual rubber or metal lamination orlayer describes a segment of a sphere and all of these layers ofdifferent radii have a common center 41 that lies upon the longitudinalaxis of symmetry 42, thereby permitting angular movement in 3 degrees offreedom about the center 41, including partial rotation of the shaft 13about said longitudinal axis, and tilting of said axis within the planeof the paper and into or out of it. The overall configuration depictedcould represent a sealed moveable shaft penetrating the hull of asubmersible vehicle for control purposes in the presence of externalhydrostatic pressure, where said spherical bearings are bearing-seals aspreviously noted.

Pusher ends 1 and 2 are each loosely anchored to and extend betweenreceptive notches that are respectively parts of loading member 7 andthe midshaft spherical loading member 4 (in broken-away sectionalviews). It will be seen that end 2 of any pusher is not constrained tomove on a lineal path, but can move over a region. Hence this negativespring means has a range of effectiveness as a negative spring about anyaxis of angular movement extending through the spherical center 41,including the axis of symmetry 42 and those either in the plane of thepaper or perpendicular to it, i.e., the two tilting axes. The tilt ofthe longitudinal axis about the point 41 through the angle aschematically indicates the latter condition, in which the surface ofspherical member 4 moves through an arched trajectory 6 relative tomember 7 resulting in incremental movements of the laminate edgesbetween them as seen by dashed lines.

It is noted that pushers with their chord lines 1-2 oriented radially inthe neutral unstrained orientation of the bearings need not be spacedequally around the spherical member 4 as indicated, but would have someeffectiveness when arranged radially around spherical member 4 in anylocation. In this spherical case, the negative spring means operates asa whole about the center 41, although the variable effectiveness of anyone pusher depends upon its orientation relative to any particular axisof angular movement.

FIGS. 12 and 13 show top and partly broken-away elevation views,respectively, of a negative-spring-compensated thrust-loaded combinationof elastomeric bearings. It includes an upper elastomeric bearing 9 astacked along a common axis 11 upon a lower elastomeric bearing 9 b,each having a circular top end plate 18 a/18 b and an annularly-extendedcircular bottom plate 19 a/19 b, with bottom plate 19 a affixed atmedian plane 15 to top plate 18 b, and each said bearing having anenveloping rotational negative spring device 16 a/16 b that is directlypaired to its associated bearing, thereby to reduce the apparent springstiffness of said bearing.

The combined elastomeric bearings support a thrust load L emplaced by aloading member 4 (dashed lines) and transmit that load through to thebase 7 (dashed lines). Limited angular movement of load member 4indicated by the arched path 6 is shared between the combined bearings,with each of the negative spring devices 16 a/16 b experiencing the sameangular rotation as its associated bearing.

Each negative spring device 16 a/16 b can be seen to be another exampleof the general representation of the FIG. 4 model. Each said deviceincludes multiple bent and nested leaf springs 3, 3 a and 3 b, eachhaving ends 1 or 1 a and 2 or 2 a as previously described, with examplesseen face-on at the cut-away side of FIG. 13 and on edge near thecenterline of same figure (bowed in opposite directions for reasons tobe explained), and seen top edge-on as radial hidden lines in the topview. Ends 1 or 1 a are loosely anchored in radially-oriented grooves ornotches in the bottom plates 19 a/19 b, and ends 2 or 2 a, are similarlyanchored in notches in top annular end plates 14 a and 14 brespectively, with the chord lines connecting said ends of each leafspring being near-axially oriented at the torque-neutral configurationof said negative spring device. Said annular end plates 14 a/14 b areconcentric with, surround and experience the same angular rotation asthe circular end plates 18 a/18 b of said elastomeric bearings, and arepreferably Annular extensions of the latter end plates, but may beseparated along the periphery of end plates 18 a/18 b though keyedtogether so that when load L is not applied by said loading member 4,the combined axial forces of said bent leaf springs are not transferredto the end plates 18 a/18 b of said elastomeric bearings 9 a/9 b tocreate undesirable tension stress in the elastomer layers. Anotherreason for separation of the negative spring and elastomeric bearing endplates may be a requirement for different heights of the negative springdevices and the elastomeric bearings to achieve their separatefunctions.

The top and bottom bearings may be considered as segments of a singlebearing, but it may be found desirable to divide the overall negativespring device into more than two segments because of the limited rangeof angular movement of any single bearing or any single tier of saidnegative spring means, seen here as leaf springs.

Another possible function or side effect of the negative springmechanism described is lateral support of the laminated elastomericthrust bearing in similar fashion to the concepts presented in U.S. Pat.No. 3,228,673. FIGS. 9 and 10 of U.S. Pat. No. 3,228,673 indicatenear-axially oriented beams that surround and span the laminate stack,being supported at their midsections by extensions of the central plateof the laminate stack and impinging along their length through rubberpads or slipper rings upon the laminate edges to support said edges.

In the current FIG. 13, the configuration of leaf springs may beengineered to provide such lateral support by acting as beams to supportand prevent unstable lateral bulging of the laminated rubber bearings,thereby permitting taller individual bearings with larger angular rangethan possible without the negative spring means.

This purpose may be extended even further to stacked bearings as shownif pairs of leaf springs 3 a and 3 b above and below the median plane 15are each effectively joined together to provide radial rigidity as ofone beam, i.e., an extended leaf spring 3 c, passing through radialslots at 1 a and 2 a in plates 19 a and 14 b. One way of doing so is tomake each said extended leaf spring 3 c wider than said radial slotsexcept for a notched segment that that is narrow enough to pass throughthe slots while still maintaining the cross-beam functionality. Anotherway would be to form each said leaf spring pair from a single metalstrip having a small through-hole provided at the point where the strippasses through said median plane, and to movably secure the strip to themedian plates with a circumferentially-oriented pin passed through saidhole and affixed to said median plates on each side of said hole.

Said effective beams could thereby provide support of the midsectionplates at the median plane 15 against lateral movement relative to theoverall top and bottom end plates 18 a/19 b. FIG. 13 shows the arc ofthe top leaf spring 3 a bent tangentially in the opposite direction fromthet of the bottom leaf spring 3 b to better enable that rigidity. Ifmore segments than two were added as discussed above, the leaf springbeam-functionality could be extended sinusoidally with continuous metalstrips along with each such segment.

In the case of serially stacked elastomeric bearings, if the reactivetorque of each were to be independently nearly perfectly compensated bythe equal and opposite negative spring torques as seen in FIG. 1, thenangular motion applied to such a stack may produce undesirable effects.Over-compensation by negative spring means could apparently result in“snap action” instability between a stacked bearing and its neighbors,while exact compensation may be indeterminate. Some under-compensationmay be necessary to insure relatively equal sharing of the totalmovement applied to the stack of bearings. Realistically, minorvariations in the compensated stiffness would tend to vary theindividual angular movements resulting from the same torque applied toall bearings in such a stack.

Therefore, the provision of connected leaf springs as discussed may alsohave a very important additional beneficial effect upon thatcompensation problem. The lateral stiffness of the carried-through leafsprings could result in an influence between neighboring bearings thatcould urge them to equalize their separate angular motions.

FIGS. 14 and 15 are top and partly-broken-away side views of amultipally-segmented laminated rubber bearing stack including laminatesegments 9, and at intervals along said stack between said segments,radially extended thicker metal layers 19 are interposed. This device issimilar to that of FIGS. 12 and 13 except that all laminate layers ofmetal and rubber (shown by slanted parallel lines) are not planar, butrather are truncated ring sections of spheres nested and bonded togethersuch that each said spherical section metal ring has an associatedcenter of revolution 41, as seen for several of the thicker layers 19.The centers of each said spherical metal ring are not shown as a singlecommon point, but are spaced apart as are the metal layers from oneanother, together forming a centerline 42 as indicated in FIG. 15.

As shown in FIG. 11, each said spherical metal ring can move slightlywith respect to its neighboring metal layers due to resilience of theintervening rubber layers in any of 3 dimensions about its center point,including angular rotation about the longitudinal axis or tilting aboutan axis within the plane of the paper and in or out of the paper plane.The gradual accumulation of such tilting of each said metal ring aboutits own center along the centerline of the stack amounts to a bending ofthe entire laminate stack as a whole, e.g., as a flexible pipe with itscenterline bent into a circular arc, the radius of which depends uponthe angle a as shown relative to the local centerline and the level ofstrain in the rubber layers. If the annulus formed between the insideand outside diameters of each ring is relatively narrow, said metallayers may be segments of a cone having a similar slope as thecross-sectional chord of the spherical-shaped rings.

Said radially extended thicker metal layers 19 are interposed to providethe supporting elements for multiple bent leaf springs 3 engaging themas was indicated for FIG. 13, except that here said bent leaf springsare shown to be formed from a continuous metal strip with its widthdimension oriented tangentially. Each strip has a small hole atintervals that is penetrated by small radial pins 44 extending from themetal layers 19 that it engages in succession and all are secured byclamp rings 45.

The bent springs formed by said strips would function within individualsegments as negative spring components as before, each pushing at itsend points 1 and 2 against its bordering extended thicker metal layers19. The bent strips have their bowed extents more or less sinusoidallybowed radially outward as seen here, movably linked as noted to eachsaid extended metal layer 19 between segments along the length of thelaminated structure. In this case, the bent springs are capable ofproviding negative spring action for the twisting degree of freedom asbefore, as well as for the tilting degrees of freedom, the latter beingdue not to the FIG. 4 model, but a consequence of the inherent negativespring property of force vs. deflection for bent columns: the columnforces F increase as the chord length of a spring between successiveextended metal layers increases on the outside of a bend of the laminatestack, urging for an increase in the bend and acting against theelastomeric reaction forces due to shear stress in the rubber. At thesame time, the column forces F will decrease on the inside of a bend asthe elastomeric reaction increases.

This device may have use as an arbitrarily elongated bearing-seal, witheach end suitably mated and sealed with respect to a cooperatingevacuated vessel element, in undersea or other external hydrostaticpressure applications, to permit negative-spring-compensated relativetilting or twisting movements between said cooperating elements. Saidpressure would cause a compression force throughout its length actingparallel to the centerline. and would cause hoop compression forces todevelop within its metal rings. Besides the radial extensions on thethicker metal rings 19, similar radial extensions 46 might be employedon the other metal rings to strengthen them, as shown for several ofthem.

FIGS. 16 and 17 show another way to use the basic mechanism of thenegative spring device of FIG. 4. These figures depict top and sidecross-sectional views of a negative spring compensation device 16 havingan inner shaft 13 and an outer cylinder 12. Circular end plates 7 arecaptured and constrained against rotation relative to cylinder 12 andare apertured to provide concentric radial bearings 19 for passage andlow-friction rotational support of shaft 13. Said shaft contains acentral circular flange or center plate 4 that is sandwiched between twocircumferentially-spaced sets of pushers 3, totaling 12 on each side inthis case. Each said pusher has ends 1 and 2 that are loosely fixed inreceptacles on the inside of an end plate 7 and on flange 4respectively. These pushers are axially oriented when the shaft 13 is inthe force-centered state of angular rotation relative to the cylinder 12and endplates 7.

It can be seen that rotation of the shaft 13 will cause the flange end 2of each pusher 3 to move along a circular path 6, its center line orchord thereby assuming a shifted position as shown by dashed lines 31 inside view FIG. 17 (for clarity, only one said pusher 3 is so indicatedon each side of flange 4). As in the simple case of FIG. 4, a componentof force in the plane of flange 4 will be generated, urging furthermovement in the same direction. These tangential forces on both sides ofthe flange, acting at the radius of the circle of pushers, create anoverall torque to on the shaft in the direction of displacement.

Functionally, if the shaft 13 and the cylinder 12, as application pointsper the FIG. 4 discussion, are respectively mechanically coupled to theopposing loading members of an elastomeric bearing and are properlycalibrated in terms of matching positive and negative spring rates withtheir torque or force centers aligned as discussed in association withFIGS. 1, 2 and 3, the purposes of the invention will be achieved.

Another way to accomplish the same effect is to make at least one of theend plates 7 axially movable toward the other and to apply an externalaxial force urging them towards each other, said axial force beingtransferred through the pushers and flange 4. This axial force could beprovided, for instance, by hydraulic or pneumatic pressure applied toend plate 7 acting as a piston in cylinder 12, or by compression springsof some type such as a large Belleville spring acting between an end ofthe cylinder 12 and the associated end plate 7. In this variation of thedevice, the pushers 3 could be equal length rigid pins or columns as inFIG. 10. Controllable axial force, as by electromagnetic or pressuremeans, would make it possible to vary the negative spring force inresponse to varying conditions of use. Using rigid pins would imply someaxial motion accompanying angular movements depending upon the cosine ofthe angle.

Alternatively, flange 4, instead of being one solid piece, could beimbued with internal axial spring properties urging its faces 4 a and 4b apart, thereby providing the same end-loading effect upon rigid pins3. Specifically, a flange face 4 a could be made compliant axially withrespect to the shaft and to the other face 4 b while still affixedangularly to said shaft and thus rotatable with it, and said two facescould be urged apart as pistons by pressure means or by force means suchas one or more Bellville springs so as to bear upon said rigid pins,resulting in the negative spring effect of the invention.

Although there is an advantage in the configuration of FIGS. 16 and 17in that the arrays of pushers 3 on both sides of the flange 4 canbalance out the high axial forces of the pushers upon it, it is obviousthat a single-sided array could also work as desired if the flange 4were provided with a suitable thrust bearing on one side to accept theaxial loading from the other side while permitting rotation of saidflange.

FIG. 18 schematically shows a side view of another related model ofnegative spring realization, including a lever arm 4 and a force source,pusher 3. Pusher 3 has a first end point 1, and lever arm 4 has length Rand a first end point 11. Both first ends are angular moveably affixedto a framework 7, while the other ends of each are jointed together atpoint 2. Movements of point 2 are constrained to follow a circular arc 6about the lever arm center 11, while the length of pusher 3 varies toaccommodate their connection. The point 5 along arc 6 where joint 2 isaligned with a line connecting end point 1 to end point 11 isgeometrically the force center, the location where the length of pusher3 is minimized and equal to Dm. A configuration of pusher 3 and leverarm 4 with their joint point 2 below point 5 is indicated by dashedlines.

Any deviation in either direction from the force center 5 along the path6 will result in a component of force fn on the joint 2 urging itfarther away from the force center 5. Force fn may thereupon be passedthrough the point 2 to an attached linkage 24 aligned roughlytangentially to path 6 at point 5. Otherwise, the tangential forceacting upon the lever arm 4 could develop a torque to upon an axisperpendicular to the paper through point 11. If either said attachedlinkage 24 or said torqued axis through point 11, as one attachmentpoint, and framework 7 as the second attachment point are eachmechanically linked respectively to one of the movable load members ofan elastomeric bearing, given the appropriate calibration in accordancewith the previous discussion of FIGS. 1, 2, and 3, the purpose of theinvention will be satisfied.

With a small angular deflection a (in degrees) of the lever arm awayfrom the force center 5, and pusher force=F, the negative torque tndeveloped on the lever arm 4 will have the approximate magnitudetn=(a/57.3)*F*R[1+(R/Dm)].

It is noted that similar comments may apply as well to the pusher 3; ifit can function as a rigid lever itself, it can experience a componentof force tangential to its rotation about its end center 1, andbeneficially transfer the resulting force or torque as a negative springdevice to compensate a linked elastomeric bearing. In fact, if thepusher involves an ordinary compression spring, the torque exerted uponit would be beneficially enhanced by its longer moment arm even as itsforce F diminishes at large excursions.

The lever arm 4 could be extended within a truncated plane in any radialdirection and radius from center point 11 to an attachment point for alinkage other than point 2 (e.g., a bellcrank configuration), with saidlinkage aligned roughly with a tangent to said radius when force centerpoint 5 is aligned with the line between points 1 and 11.

In a variation similar in principle to those discussed previouslyrelative to FIG. 10 and variants of the device of FIGS. 16 and 17, theend point 1 of the pusher 3 could be made movable with respect to frame7 and urged by a force towards point 11. This would again permit thepusher 3 to be made inextensible while still developing an increasingtorque upon lever arm 4 with increasing angle, as desired.

To this point, FIG. 18 has been treated as a two-dimensionalrepresentation, but if the pusher 3 and lever arm 4 are not constrainedto up/down movement in the plane of the paper, but rather are freelypermitted any three-dimensional path in/out of the paper as well, anegative spring effect could be provided for a multiaxis sphericalelastomeric bearing as seen in FIG. 11, said bearing's center point 41itself serving as the current joint point 11, with said similarly freepusher jointed (as with a ball-and-socket joint) to said sphericalbearing's longitudinal shaft at a point 2 in line with point 1 and point11 of said pusher in the neutral position, thereby to reduce stiffnessin said spherical bearing's tilting axes. This is, in effect, what oneof the small pushers 3 in FIG. 11 does with respect to its lever armfrom point 2 to center point 41.

As seen, the pusher 3 and the lever arm 4 face each other and both endpoints 1 and 11 and arc 6 (including point 5) are represented as beingin the vertical plane of the paper. The location of the end point 1 ofpusher 3 need not be in the same vertical plane as lever arm 4 and itsarc 6. Indeed, in 3 dimensions, with the left-facing lever arm 4 asshown having an axis through point 11 extending perpendicular to thepaper, the negative spring effect will exist with locations of the endpoint 1 of the pusher out of the plane of the paper and to the left ofsaid axis. In this case, the force center point 5 on the arc 6 will bedefined by the arc's intersection with a plane that contains end point 1and said axis through point 11.

FIG. 19 shows the plan view of two pushers 3 linked to the lever arm 4,arranged on either side of the lever arm which moves on a (notillustrated) arched path within a vertical plane perpendicular to thepaper. It will be obvious that equal force pushers could balance out thehorizontal component of each other's force, but may need ball-and-socketend joints to follow the curvature of the lever arm's path.

It is also noted that the FIG. 19 configuration can be recognized asanother mechanical model for the cylindrical negative spring device ofFIGS. 16 and 17, where any pair of the pushers 3 on opposite sides ofthe flange 4 create a combined tangential force on the lever armrepresented by their radius from the center of the shaft 13, and theyare paralleled in action with all the other pairs to create the overalltorque vs. deflection property of that device.

FIG. 20 shows a configuration of the same components as in FIG. 18,except that the force source 3 is a tension device pulling upon itsconnection to the end joint 2 of the lever arm 4, i.e., it is a “puller”such as a pull spring. This tension device 3 is arranged so that its endpoint 1 is on the opposite side of end point 11 of the lever arm 4 fromthe direction that lever arm 4 extends, so that it still creates acompressive force within the lever arm as does a pusher. Again, there isa point 5 on the arc 6 of the lever arm that is aligned with anextension of the line from end 1 to end 11, but this is geometricallythe point at which the puller length is maximum, and is the forcecenter. The position of the lever arm and tension device when theirjoint 2 is below point 5 is indicated by dashed lines.

As before, it can be seen that the “over-center” characteristic applies:any angular deviation a of the lever arm from the force center 5 willresult in a tangential force developed on the end of the lever arm thaturges further deflection in the same direction, i.e., the requirement ismet for a negative spring, given that the lever arm 4 and frame 7 arefunctionally linked to elastomeric bearing means via attachment pointsas previously explained. With puller force=F, lever arm length=R and thedistance between end points 1 and 11=DI, the negative torque tndeveloped has the approximate magnitude for small values of angle a,tn=(a/57.3)*F*R*[1/(1+(R/DI))].

A puller device would generally require a stronger force source than thepusher type to get the same results with similar dimensions, as can bededuced from examination of the two corresponding force equations.

Regarding 3-dimensional variations, similar considerations exist in thiscase as with pusher force sources, except that the end point 1 of thepuller 3 must be on the side of the lever axis through point 11 that isopposite from the direction that lever arm 4 extends. Furthermore,analogously to the use of multiple pushers as exemplified by FIG. 19,multiple pullers having end points opposite the direction of lever armextension from its axis 11 could be applied to the lever arm 4 to obtainthe required negative spring effects.

There is a significant application of these principles for use inelastomeric bearing-equipped helicopters, as well as others that havetheir rotor blades retained by tension-torsion straps that also havetorsional spring properties. When laminated elastomeric bearings areused for blade retention on a helicopter, they create a torsional springeffect upon the pitch control linkages as the pitch of the rotor bladesis changed from an average neutral position at which the elastomericbearings are relaxed, i.e., untorqued. With helicopters that lack aforce boost system, these control forces react against and must normallybe borne directly by the pilot's hands on the controls, and in somecases may result in forces that are excessive.

The pilot's controls include a collective pitch stick, which moves allof the blade angles together, through a so-called swash plate, to changethe overall lift of the rotor, while in order to obtain directionalcontrol, the cyclic pitch stick causes the individual blades to bevaried up and down sinusoidally through a small amplitude relative tothe average pitch set by the collective system, once per revolution ofthe rotor. As noted, the forces reflected as a result of angularmovement of said elastomeric bearings may exceed the pilot'scapabilities for direct control, especially of the collective pitchstick which produces a greater range of blade pitch angle than thecyclic stick.

Many of the negative spring devices discussed herein could be used tocompensate for these forces. One way to realize that compensation wouldbe to apply a torsional negative spring device within the rotor hubitself to each individual blade's elastomeric bearing directly. Thatcould be effective in reducing feedback forces, but would have thedisadvantages of requiring rapid oscillatory motion with consequent wearof the negative spring devices, while increasing rotor hub bulk andcomplexity.

Rather, since the average per cycle elastic reactive forces on thecollective pitch control linkage would have the same kind offorce-deflection properties as those described for individualelastomeric bearings, a negative spring device could be applied withinthat collective control linkage, separated from but effective for theindividual elastomeric bearings. Said negative spring device wouldcreate forces that act in the direction of movement of the collectivestick from the neutral position, opposite to the forces reacting againstsaid stick movement that result from reaction torques of all theelastomeric retention bearings moving simultaneously. By these means,only the feedback force of the collective part of the angular deflectionof all the rotor blade retention elastomeric bearings together would becompensated when said negative spring device was properly calibrated asdiscussed.

Feedback forces on the cyclic pitch stick may also be reduced. Theseforces include blade pitch axis inertial reaction forces as well as thespring forces of the elastomeric retention bearings. Beside the possibleuse of force boosters, there is the possibility of reducing reactionforces by balancing the inertia and spring effects against each other,as pointed out in Canadian Patent 731007 (Ballauer, pg. 13-17). Thatmethod has to do with the oscillatory interchange of potential energy ofthe spring into kinetic energy of the blade's pitch angle velocity andvice-versa during each rotor rotation cycle, and is maximized in effectwhen the resonant frequency of the mass-spring system defined by theblade pitch-axis moment of inertia and the elastomeric spring constantplus some aerodynamic reactive torque is the same as the rotor RPM.Proper use of this technique can greatly reduce cyclic stick feedbackforces.

FIG. 21 shows a side view of the collective pitch control stick 20 of ahelicopter, pivoted at point 11 near the floor of the cabin to a framemember 7. The range of movement is from the full down position 21 to thefull up position 23 (both these extremes being shown by dashedcenterlines), which thereby defines the range of collective pitch of thelinkage-attached rotor blades and their associated retentive elastomericbearings. An intermediate neutral position 22 may correspond to therelaxed, untorqued angular orientation of said bearings. The collectivepitch control stick may typically be extended rearward of its pivotpoint 11 as shown to provide the lever arm 4 which links at joint 2through a control rod 24 to the collective pitch mechanism in order toperform its primary mission.

In compliance with the intent of this invention, the rearward extensionlever 4 may be employed for force counteracting purposes as in FIG. 18.Also in accordance with FIG. 18, a compression spring or another type ofpusher 3 is linked to the lever arm 4, conveniently also at joint 2, andis rotatably joined to the frame at its end point 1.

Alignment of lever arm 4 with the line between the points 1 and 11defines the center or neutral force position of the negative springmeans, corresponding to the stick's position 22 as shown. In saidneutral position, the force vector of the compression spring 3 will bein line with the pivot point 11, and no torque can be exerted by thepusher 3 upon the collective pitch stick. But an upward movement of thestick handle will result in an upward force fn upon the handle from thepusher acting upon the rearward extension 4, and a downward force willresult when the stick is moved downward from its neutral position. Thegreater the movement in either direction, the greater the force in thesame direction, thereby cancelling the force fe caused by the resilienttorque reaction of the combined blade pitch elastomeric bearings asreflected back to the collective stick over the effective range.

This would remove much or most of the overall force felt by the pilot,assuming that the intent was to align the negative spring force centerwith the elastomeric neutral point, or otherwise, as discussed withrespect to FIG. 3. As pointed out there, by offsetting the force centerpoint of the negative spring mechanism relative to the untorqued orrelaxed position of the elastomeric bearings, it would be possible tocreate an always up or always down combined force on the stick, whichmight be desired to assure that hands-off operation would move the stickto the safest condition. Otherwise, such offsets or incompletecompensations may be useful in some helicopters for counteractingaerodynamic forces on the blades.

It will be obvious based on prior discussion that instead of making useof the rearward extension of the stick to provide the lever arm 4 asshown, that the negative spring assembly (including base 7, pusher 3,and lever arm 4) could be located via its pair of attachment points inany appropriate position along the linkage system between the collectivepitch stick and the swash plate, given that the path 6 is sufficientlyaligned with the movement of the local control linkage as it is withcontrol rod 24.

As discussed relative to FIG. 20, another means of realizing a negativespring mechanism is to use a pull source instead of a pusher acting uponthe collective stick. In a very simple example, a pull spring could beattached on one end to a point on the collective pitch stick part of theway from its pivot 11 to its handle, and on the other end to a framepoint some distance behind that center pivot and in the planeperpendicular to the paper containing points 2 and 11 and the neutralpoint 5.

FIG. 22 indicates a collective pitch stick 20 with an attached negativespring cancellation device 16 according to FIGS. 16 and 17. The cylinderhousing 12 and shaft 13 of device 16 are seen end-on in this side view.The shaft 13 is linked to or provides the axis that extends throughpoint 11 perpendicular to the paper and is affixed to the collectivestick 20 so that said shaft shares in the same or proportional angulardeflections of the stick and transfers its negative torques to the stickand its extended lever arm 4. Cylinder 12 is enclosed in a circularlyadjustable clamping bracket 8 that is affixed to the frame 7. Rotationof cylinder 12 within the clamp 8 before tightening permits alignment ofthe torque center of the negative spring device with respect to thecombined elastomeric bearings through the swash plate as desired forcalibration as previously discussed.

Given that the collective pitch stick experiences the reflected averagespring force developed by the combined elastomeric bearings as they aremoved through a part of their angular range, the negative spring rate ofthe cancellation device 16 is designed to match at the stick, more orless, the positive spring rate of said combined elastomeric bearings.

The foregoing description shows various instances of the pairing ofnegative spring devices with elastomeric bearings. “Pairing” refers tothe fact that an elastomeric bearing separates two opposed relativelylaterally moveable loading members that develop reactive forces ortorques between them, and each of two attachment points of an associatednegative spring device is mechanically connected respectively with oneof said two members, such that said negative spring device moves atleast partly with said elastomeric bearing or bearings to developcompensating or negating lateral forces or torques.

FIGS. 7-9 and 11-15 show instances of direct side-by-side pairing of anelastomeric thrust bearing with a negative spring device, while theremainder of the figures may be remote-coupled or partly-coupled withone or more elastomeric bearings. When directly coupled, both theelastomeric thrust bearing and the associated negative spring devicegenerally share the same lineal or angular ranges, whereas in theremote- or part-coupled cases, the shared movements may be merelyproportional or even nonlinearly related.

Upon consideration of the variety of negative spring devices consideredhere, it is observed that there are unifying properties of them all:They all involve linking mechanical elements, i.e., lever arms (that maybe pushers as well) or pins, that experience compressive force betweentheir opposed ends and whose ends are moveably fastened to supportingelements that are subjected to relative lateral movements. The resultingdiagonality of the linking elements transfers a component of saidcompressive force onto their supporting elements in the direction oftheir relative lateral movements. The two supporting elements of thenegative spring means are respectively mechanically coupled togetherwith the two opposed loading members of at least one elastomeric bearingor seal such that movement imparted to the negative spring means isexperienced, at least in part, as movement of said at least one bearingor seal whose combined positive spring rate is at least partiallycompensated by the negative spring rate of said negative spring devices.

It will be understood that the embodiments described above are merelyexemplary and that persons skilled in the art may make many variationsand modifications without departing from the spirit and scope of theinvention as defined in the appended claims.

I claim:
 1. A force compensation system comprising: at least oneelastomeric or laminated rubber bearing having two opposed loadingmembers, a positive spring rate and a center or neutral position; andnegative spring means having a negative spring rate and a center orneutral position, comprising: at least one elongated member or pusherhaving two opposed ends, and a source of force imposing compressiveforce between said ends of each of said elongated members, and tworeceiving and supporting members that moveably support and respectivelyeach receive force applied by one of said two ends and are constrainedin lateral motion relative to each other, whereby a component of saidcompressive force will be developed upon said receiving and supportingmembers in a direction of said constrained lateral motion; and said tworeceiving and supporting members of said negative spring means arerespectively mechanically coupled together with said two opposed loadingmembers of said at least one elastomeric bearing such that any saidrelative lateral movement imparted to said receiving and supportingmembers of said negative spring means is experienced substantiallyproportionately as movement between said two opposed loading members ofsaid at least one elastomeric bearing, whereby any change in net torquedeveloped upon said elastomeric bearings accountable to their saidpositive spring rate is at least partially compensated by the change iinet torque developed by said negative spring rate of said negativespring means.
 2. The force compensation system of claim 1 wherein saidtwo receiving and supporting members of said negative spring means arerespectively directly coupled together with said two opposed loadingmembers of said at least one elastomeric bearing.
 3. The compensationsystem of claim 2 wherein a multiplicity of elastomeric bearing segmentsare stacked to define an overall elastomeric bearing with extendedlayers defining the opposing loading members of each said segment, and aplurality of compression springs are arrayed such that each of theirsaid ends bears respectively upon one of said extended layers, therebyserving as negative spring means to compensate for the positive springrate of said bearing segment between said extended layers.
 4. The forcecompensation system of claim 3 wherein at least one said compressionspring serving a said bearing segment is functionally joined as a beamwith a said compression spring serving an adjacent said bearing segment,whereby said joined beams mutually affect the extent of motionexperienced by each said adjacent segment.
 5. The force compensationsystem of claim 4 wherein said compression springs are leaf springs withtheir edges aligned radially with respect to an axis.
 6. The forcecompensation system of claim 4 wherein said compression springs are leafsprings with their edges aligned circumferentially with respect to anaxis.
 7. The force compensation system of claim 2, wherein saidelastomeric bearing has laminations that are cylindrically disposedabout an axis and said negative spring means comprise a multiplicity ofcompression springs.
 8. The force compensation system of claim 2,wherein said elastomeric bearing has laminations that are sphericallydisposed about a center and said negative spring means comprise amultiplicity of compression springs.
 9. The force compensation system ofclaim 2, wherein said elastomeric bearing has laminations that eachencircle a center point and said negative spring means comprise amultiplicity of compression springs.
 10. The force compensation systemof claim 1 wherein a said receiving and supporting member includes atleast one lever arm having a point of application of force from an endof said at least one elongated member and a center point about which itis rotatable, whereby a component of tangential force is created uponsaid point of application.
 11. The force compensation system of claim 10wherein said receiving and supporting member includes a member rotatablethrough an angle about an axis and a multiplicity of said elongatedmembers are arrayed circumferentially with their said ends bearing uponsaid rotatable member at radii acting as said lever arms.
 12. The forcecompensation system of claim 1 wherein said negative spring rate issmaller in magnitude than said positive spring rate, whereby said nettorque of said elastomeric bearings is under-compensated.
 13. The forcecompensation system of claim 1 wherein said negative spring rate islarger in magnitude than said positive spring rate, whereby saidnet-torque of said elastomeric bearings is over-compensated.
 14. Theforce compensation system of claim 1 wherein said negative spring rateis adjustable in magnitude.
 15. The force compensation system of claim 1wherein said center positions of said elastomeric bearings and saidnegative spring means do not coincide.
 16. The force compensation systemof claim 1 wherein said center position of said negative spring means isadjustable.
 17. The force compensation system of claim 1 wherein saidsource of force is at least one compression spring or pusher.
 18. Theforce compensation system of claim 1 wherein said source of force is atleast one leaf spring.
 19. The force compensation system of claim 1wherein said source of force is at least one pull spring.
 20. The forcecompensation system of claim 1 wherein said source of force is a hoopacting as a tension or compression spring that is a said receiving andsupporting member and imposes compressive forces between said ends ofsaid elongated members that are pins.